HVAC

Tools in this group

• Manual J Cooling Load (Simplified) - Simplified sensible and latent cooling load estimate.
• Manual J Heating Load (Simplified) - Simplified heating load estimate.
• Duct Sizing - Round and rectangular duct size from CFM and friction rate.
• Static Pressure - Total external static pressure from duct elements.
• Refrigerant P-T Chart - Pressure-temperature for common refrigerants.
• Superheat and Subcool - Superheat or subcool from line temperature and pressure.
• SEER and EER Conversion - Bidirectional rating conversion.
• Heat Pump Balance Point - Balance temperature from heating capacity and heat loss.
• Sensible Heat Ratio - SHR from total and sensible cooling loads.
• CFM per Ton - Standard 400 CFM per ton with humidity adjustments.
• Combustion Air - Required combustion air opening from BTU input and room volume.
• Compare Two Refrigerants - Side-by-side P-T at a chosen pressure or temperature with manufacturer attribution.
• Refrigerant Charge Weighing - Charge in oz from line-set length, diameter, and refrigerant.
• Approach and Delta-T Diagnostics - Condenser approach and supply/return delta-T with band labels.
• Outdoor Air Mix - Mixed dry-bulb and humidity ratio from RA/OA states and OA fraction.
• Equivalent Length of Fittings - Total equivalent feet from fitting type and diameter counts.
• Wet-Bulb Sling Psychrometer - RH, dew point, and GPP from dry-bulb and wet-bulb.
• Pipe Insulation Thickness - Required thickness from cylindrical conduction at a target surface temp.
• Latent Heat Evaporative Cooling - Cooling effect from evaporation rate and h_fg.
• Fan Affinity Laws - Q ~ N, P ~ N^2, kW ~ N^3 across RPM, CFM, SP, and kW changes.
• Belt Length and Pulley Speed - V-belt length, driven RPM, and belt speed from pulley diameters and centers.
• Compressed Air Receiver Sizing - Required receiver gallons from tool CFM-at-duty-cycle list and pump.
• Geothermal Loop Length - Estimated loop length from design BTU and BTU-per-foot benchmarks (simplified).
• Hydronic Baseboard Output - BTU/hr from manufacturer-attributed BTU/ft tables interpolated by water temp.
• Pump NPSH Available - Atmospheric, vapor, static, and friction heads to NPSHa with cavitation flag.
• Duct Leakage Test-and-Balance - SMACNA leakage class from measured-vs-design CFM at test pressure; pass/fail against the design class.
• Residential Duct Leakage CFM25 (IECC R403.3.5) - Whether a residential duct system passes the energy-code tightness test: the measured total leakage at 25 Pa (from a duct blaster) normalized to the conditioned floor area, CFM25 per 100 ft^2 = leakage / area x 100, against the IECC limit (4 total / post-construction, 3 for a rough-in without the air handler). 80 CFM25 on 2,000 ft^2 gives exactly 4.0 -- a pass at the limit; a tighter 60 CFM25 gives 3.0, a leakier 100 gives 5.0 (fail). Distinct from the SMACNA leakage-class test for commercial duct. A field aid; the adopted energy code, the required test type, and the rater govern.
• ASHRAE 62.1 Outdoor-Air Ventilation - Breathing-zone and zone outdoor airflow per ASHRAE 62.1 §6.2.2.1. User supplies Rp / Ra from the AHJ-adopted edition; the tile does not bundle Table 6-1.
• Commercial Kitchen Hood Exhaust (IMC 507) - Type I (grease) and Type II (vapor-only) commercial-kitchen hood exhaust airflow per IMC 2021 §507.13 / §507.20 duty multipliers. Makeup air, duct area, grease-duct slope reminder. NFPA 96 governs grease-handling exhaust system design.
• Sensible Heat Ratio / Latent Split (ASHRAE) - Field tile: given total cooling Btu/hr, return-air dry / wet-bulb, supply dry-bulb, CFM, and altitude, returns sensible / latent split (1.08 / 4840 coefficients with altitude correction), SHR, and return / supply humidity ratios (grains/lb). Companion to the shr tile which takes both Q values directly. Per ASHRAE Fundamentals 2021 Ch. 1 / Ch. 18.
• Duct Friction Loss and Static Pressure - Darcy-Weisbach with Swamee-Jain Colebrook; round or rectangular; fitting losses from a C_o library; total external static.
• Refrigerant Superheat / Subcooling (psig-psia toggle) - Saturation T from bundled P-T tables (R-410A / R-32 / R-454B / R-22 / R-134a); per-input psig vs psia toggle; in-range / low / high flags.
• Cooling Tower Approach and Range - Range, approach, heat rejection (gpm × 500 × range BTU/hr), and fan kW per ton against typical 5-10 °F approach / 8-12 °F range targets.
• Pipe Insulation Heat Loss (bare vs insulated) - Cylindrical conduction + convection + radiation; effectiveness % vs bare; six insulation k-values with manufacturer attribution.
• Chiller Tonnage (Delta-T and GPM) - Cooling capacity (tons, BTU/hr, kW) from chilled-water flow and the entering/leaving temperature split; required flow at a nameplate tonnage; water and 30/50% propylene-glycol factors. Q = GPM × 500 × delta-T (water); fluid factors per ASHRAE Fundamentals 2021 Ch. 31.
• Heat Exchanger LMTD and Effectiveness-NTU - Log-mean temperature difference, heat duty, required UA, effectiveness, NTU, and capacity-rate ratio for counter-flow or parallel-flow from four temperatures and the two flows. Per the TEMA standards and standard heat-transfer texts.
• Air Changes per Hour (ACH) - ACH = CFM × 60 / volume, net delivered ACH and pressurization airflow from unequal supply/return, and a comparison against ASHRAE 62.1 / 170 occupancy targets (residential to operating room).
• Boiler Distribution Pipe Sizing - Hydronic GPM = Q / (500 × delta-T), the smallest copper / steel / PEX size at or below the velocity ceiling, velocity, Hazen-Williams head loss per 100 ft, and pump head at the run length. Per ASHRAE Systems and Equipment 2020 Ch. 13.
• Compressor Short-Cycle Protection - Estimated cycles per hour from the ASHRAE/AHRI part-load cycling parabola, on/off time, and the minimum oil-return runtime and pressure-equalization delay by system type, flagging oversized single-stage short-cycling. Per Copeland AE Bulletin 17-1226.
• Humidifier Capacity (RH Target) - Moisture addition (lb/hr, gal/day) and latent load from supply CFM and the entering-to-target RH rise, with altitude-corrected humidity ratios. Per ASHRAE Fundamentals 2021 Ch. 1 psychrometrics.
• Filter Pressure Drop and Fan-Energy Penalty - Clean and change-out pressure drop by MERV / HEPA class (velocity-scaled, user-overridable cut-sheet defaults), the fan power at each via brake-HP, and the annual fan energy and loading penalty over a clean filter. Per ASHRAE 52.2-2017 and first-principles fan power.
• Window Solar Heat Gain and Conduction Cooling Load - The two ways a window adds to a cooling load: the solar radiation through the glass (area x SHGC x peak solar factor) and the conduction driven by the indoor-outdoor difference (area x U x CLTD). On a west wall in summer the solar term dwarfs the conduction by an order of magnitude, which is why orientation and SHGC, not U-value, drive a glass cooling load - the number to diagnose a hot room or justify a shading or low-SHGC retrofit. The peak solar factor is read from the ASHRAE/ACCA table for the orientation and latitude. One Manual J component, not the stamped load sheet.
• Internal Heat Gains: People, Lighting, Equipment - The heat the building makes inside: occupant sensible and latent gain from the activity-level rate, plus lighting and plug load at 3.412 Btu/h per watt, scaled by the use factor actually on. The sensible/latent split sets the coil's job - a packed conference room flips from sensible-heavy to latent-heavy, and that latent is moisture a sensible-only 'more airflow' fix never removes. The per-person rates come from the ASHRAE activity table. One Manual J component, not the stamped load sheet.
• Opaque-Envelope Conduction Cooling Load (Sol-Air CLTD) - The conduction cooling load through a sunlit wall or roof: U x area x the sol-air CLTD, which for a dark roof in summer can run 70 F or more against a 95 F design day because the sun heats the surface well above the air. Dropping the surface's solar absorptance with a reflective cool-roof membrane cuts the load far more than the air temperature would suggest - the entire case for a cool roof in one comparison. The U-factor is the whole-assembly value; the CLTD is from the ASHRAE/ACCA table. One Manual J component, not the stamped load sheet.
• Manual D Friction Rate (Available Static Pressure) - ASP = blower ESP - component drops, friction rate FR = ASP x 100 / total effective length (in wg/100 ft). 0.60 ESP, 0.42 drops, 180 ft TEL -> 0.18 ASP, 0.10 FR (a healthy target); a 300 ft run softens it to 0.06, needing larger ducts. Drops above the blower static flag an unworkable system. The full Manual D layout governs.
• ERV Total Enthalpy Recovery - Q = 4.5 x CFM x effectiveness x (h_oa - h_ra), supply enthalpy = h_oa - eff x (h_oa - h_ra). 1000 CFM, 0.75, 38 -> 28 Btu/lb -> 33,750 Btu/hr recovered, 30.5 Btu/lb supply; a winter 12 vs 28 Btu/lb recovers -54,000 Btu/hr (heating). Feed enthalpies from moist-air-enthalpy. The ERV's rated effectiveness governs.
• Radiant Floor Heat Output - q = 2 x (T_surface - T_room)^1.1 Btu/hr-ft^2, with the ~85 F surface comfort cap (~39 Btu/hr-ft^2 in a 70 F room). An 85 F floor over a 70 F room gives 39.3; solving inverse, a 30 Btu/hr-ft^2 load needs an 81.7 F surface. Solves either direction. The panel manufacturer's ratings and radiant-loop-sizing govern.
• Hydronic Snowmelt Load and Boiler Sizing (ASHRAE) - The heated-driveway sizing balance: warm the falling snow to the 33 F film, melt it (746 Btu/hr-ft^2 per in/hr of water equivalent), and over the snow-free fraction pay the wet surface's wind-driven convective and evaporative losses - q_o = q_s + q_m + A_r(q_h + q_e), the ASHRAE / Chapman steady-state flux. The class is the design decision: the same 20 F, 10 mph storm runs 78 Btu/hr-ft^2 on a residential drive allowed to whiten (A_r = 0), 111 on a commercial walk (0.5), 144 on a must-stay-black hospital ramp (1.0). Boiler = flux x area plus ~20% back loss. A sizing aid for the chosen storm, not an annual energy or a stamped design.
• Economizer Enthalpy Changeover - Differential-enthalpy: enable free cooling when outdoor enthalpy < return (accounts for humidity); fixed-dry-bulb: enable below a setpoint. OA 24 vs RA 28 Btu/lb -> enable (4 margin); a humid OA 32 vs 28 locks out even if cool; fixed 70 F vs 65 F setpoint locks out. The ASHRAE 90.1 high-limit governs.
• Cooling Coil Face Velocity and Carryover Check - Face velocity = CFM / coil face area, checked against the ~500 fpm wet-coil carryover limit. 2000 CFM on a 24x18 coil (3.0 ft^2) -> 667 fpm, carryover risk; 1200 CFM -> 400 fpm, safe. Above the limit a wet coil blows condensate past the drain pan. The coil manufacturer's rated face velocity governs.
• VAV Box Minimum and Maximum Airflow - Max = zone sensible / (1.08 x supply deltaT), min = max(ASHRAE 62.1 ventilation, turndown x max). 12,000 Btu/hr, 20 F, 100 CFM vent, 0.30 turndown -> 556 max, 167 min (turndown governs); a denser zone at 250 CFM vent is driven to 250 by fresh air. The box range and ventilation calc govern.
• Duct Velocity Pressure - Air velocity from velocity pressure or the inverse via V = 4005 x sqrt(VP) for standard air. Per ACCA Manual D / ASHRAE Fundamentals.
• Refrigerant Line Velocity and Oil Return - Line velocity from mass flow, inside diameter, and specific volume, with an oil-return verdict against the riser / horizontal minimum and the ~4000 fpm noise ceiling. Per the ASHRAE Refrigeration Handbook.
• Air-Side Economizer Free-Cooling Hours - Sensible free-cooling capacity (BTU/hr) and annual ton-hours of mechanical cooling offset from supply airflow and the mix-to-supply delta-T over the economizer-eligible hours.
• Insulated Pipe Heat Loss (Radial) - Radial (cylindrical) conduction heat loss per linear foot and total through pipe insulation via the log-mean form, from outer diameter, insulation thickness and k-value, and the temperature difference.
• Fan Brake Horsepower - Air horsepower, brake horsepower, and the next standard NEMA motor size from airflow, total static pressure, and fan/drive efficiencies.
• Excess Air from Flue-Gas O2 - Turn a combustion analyzer's oxygen (or CO2) reading into percent excess air: EA = O2 / (20.9 - O2) x 100, or from carbon dioxide EA = (CO2max / CO2 - 1) x 100 (CO2max ~ 11.7% gas, 13.7% propane, 15.3% oil). The O2 form assumes complete combustion, so measurable CO understates it; a gas appliance targets ~3-4% O2 (15-25% excess air). A tuning aid, not a certified combustion test.
• Air-Free CO Correction - Correct a combustion analyzer's carbon monoxide to an air-free basis: CO_air_free = CO_measured x 20.9 / (20.9 - O2). As-measured CO is diluted by excess and dilution air and reads deceptively low, so a dangerous appliance can look acceptable; the ANSI Z21 400 ppm limit is air-free. Sample in the flue before the draft hood. A safety-screening aid, not a certified combustion test.
• Draft-Hood Dilution Ratio - How much room air the draft hood pulled in, from two O2 readings: dilution_ratio = (20.9 - O2_appliance) / (20.9 - O2_diluted), and the dilution-air fraction (20.9% ambient O2 mass balance). A clean appliance at 5% O2 that reads 12% O2 at the vent has pulled in 44% dilution air - so a CO reading there is only 56% of the true flue value, which is why the air-free CO sample must be taken before the draft hood. A diagnostic aid, not a certified combustion test.
• Theoretical Chimney Draft - Natural draft from the hot-vs-cold density difference: D_t = 0.52 x B x H x (1/T_o - 1/T_m) with temperatures absolute (Rankine), in inches of water column, plus a net-available estimate (factor 0.5-0.8). This is the no-flow draft; the barometric pressure must be altitude-corrected (thinner air cuts the draft) and T_m is the mean flue temp. A design aid, not a venting sign-off.
• Flue-Gas Combustion Efficiency (Stack Loss) - The third number on the analyzer screen, from the same O2 and stack-temperature readings: flue CO2 = CO2max x (1 - O2 / 20.9), dry stack loss qA = dT_C x (A1 / CO2 + B) (Siegert, per-fuel A1/B), efficiency = 100 - qA on the net (LHV) basis, plus an approximate gross (HHV) conversion because US analyzers display HHV - a gas furnace at 5% O2 and a 400 F stack over 70 F air is 90.7% net but about 81.8% on a US analyzer, right where a healthy non-condensing furnace reads. Sample dry in the undiluted flue; a condensing appliance recovers latent heat this counts as lost. A tuning aid, not a certified combustion test or an AFUE rating.
• Combustion Lambda and Air-Fuel Ratio - The lambda and air-fuel ratio a modern analyzer shows beside excess air, from the same flue oxygen: lambda = 20.9 / (20.9 - O2), excess air = (lambda - 1) x 100, AFR_actual = lambda x AFR_stoich (17.2 natural gas, 15.5 propane, 14.5 #2 oil, by mass). A gas appliance at 3% O2 is lambda 1.17, 16.8% excess air, 20.1:1 AFR against the 17.2:1 stoichiometric - the same tune three ways. Lambda = 1 + excess_air/100, the number the instrument happens to display. A tuning aid, not a certified combustion test.
• Liquid-Line Subcooling to Prevent Flash Gas - How much subcooling a system needs to keep the liquid line from flashing: dP_lift = 0.43 x lift (R-410A liquid), dP_total = dP_lift + friction, required_subcool = dP_total / P-T slope (~5 psi/degF R-410A). Techs credit only the friction and forget the vertical-lift static column that dominates on a tall riser. A design aid, not a commissioning measurement.
• Hydronic Buffer Tank Sizing (Anti-Short-Cycle) - Buffer-tank volume to stop a boiler or chiller short-cycling: V = on_time x (source_min - zone_load) / (500 x delta_T). The worst case is at nearly zero load, when the full minimum output has nowhere to go but the tank -- sizing at the design load badly undersizes it. The same formula sizes a chiller buffer. A sizing aid, not the manufacturer's data.
• Buffer Tank with Distribution-Loop Credit - Net buffer tank after crediting the water already in the distribution loop: gross V = on_time x (source_min - zone_load) / (500 x delta_T), loop water = 0.0408 x d^2 x L gal, net = max(0, gross - loop). A fat, long common primary loop already holds gallons that count toward the anti-short-cycle volume, so the tank you actually buy is smaller -- or unnecessary. A sizing aid, not the manufacturer's data.
• Round-to-Rectangular Duct Equivalent - ASHRAE equal-friction equivalent diameter between round and rectangular duct, with an aspect-ratio flag.
• Total External Static Pressure - Total external static pressure summed from a component drop list (filter, coil, registers, duct runs) against the blower fan-table rating, with a pass/fail (ACCA Manual D).
• Refrigeration Compression Ratio - Absolute compression ratio from suction and discharge gauge pressures (with altitude correction) and a high-ratio flag for single-stage limits (ASHRAE).
• Wall Assembly R-Value - The whole-assembly R-value of a framed wall by the parallel-path method, where the studs short-circuit the cavity insulation -- the wall performs below its center-of-cavity R. Reports both.
• Blown Insulation Coverage - The bags of blown insulation an attic needs at a target R, from the area and the manufacturer's bags-per-1,000-sq-ft figure, with the minimum installed (settled) thickness.
• Heat-Pump Seasonal Heating Energy and Cost vs Gas and Resistance - The number that drives every electrification job: what a heat pump costs to run for a heating season, set beside the same load delivered by a gas furnace and by electric resistance. The seasonal load (MMBtu) divided by the HSPF (AHRI 210/240 Btu per Wh) priced at the electric rate, versus load / AFUE at the gas rate and resistance at COP 1. The same heat pump is a bargain in one utility territory and a premium in another, so the answer is a calculation from the local rates, not a slogan. HSPF is the rated regional value and the load comes from a Manual J. An operating-cost estimate, not a metered bill.
• Dual-Fuel Economic Switchover (Heat Pump vs Gas Balance Point) - The setpoint a dual-fuel (hybrid) thermostat should hand off to gas at: as the outdoor temperature falls the heat pump's COP drops until the gas furnace becomes cheaper per delivered Btu. Heat-pump $/MMBtu = 293.07 / COP x rate_kwh against gas $/MMBtu = 10 / AFUE x rate_therm gives the crossover COP; the outdoor temperature where the unit falls to it is the economic switchover. Set it too warm and the customer burns gas the heat pump could deliver cheaper; too cold and they run the heat pump into its expensive low-COP range. An economic setpoint aid, not a controls-commissioning procedure.
• Heat-Pump Cold-Temperature Capacity and Auxiliary Heat - Whether a heat-pump conversion holds on the design day: a heat pump's capacity falls as it gets colder, so interpolating the AHRI 47 F and 17 F rating points down to the design temperature gives the delivered capacity, and the shortfall below the Manual J design load is the auxiliary strip heat (kW) the installer must add. A unit sized on its 47 F nameplate is badly undersized at 5 F - the classic cold-climate conversion failure. Prefer a published low-temperature data point over extrapolation when a conversion hinges on it. A sizing check, not a performance guarantee.
• Compressed-Air Leak Load and Annual Cost (Load/Unload Test) - Turns a stopwatch into a repair budget: with production off, the fraction of a cycle the compressor runs loaded (t_load / (t_load + t_unload), the DOE Compressed Air Challenge method) times the rated capacity is the leakage flow, converted to kilowatts at the system specific power (18-22 kW per 100 cfm) and to dollars at the run hours and rate. A neglected system loses 20-30 percent of its air to leaks, a well-run one under 10, and the difference is thousands a year on the plant's most expensive utility. Needs no flow meter. An estimate from a stopwatch test, not a metered audit.
• Compressed-Air Compression Power (Isentropic) and Energy Cost - The horsepower and running cost to make a given airflow, from first principles: the single-stage adiabatic (isentropic) work hp = 0.004364 x P1 x Q x (k/(k-1)) x [(P2/P1)^((k-1)/k) - 1] with k = 1.4, divided by the overall wire-to-air efficiency and priced over the run hours. It tells a tech whether a 100 cfm demand needs a 20 or 25 hp compressor and how much more a higher discharge pressure draws (100 cfm at 100 psig is ~18 hp; the same at 125 psig climbs 13 percent). Multi-staging with intercooling beats single stage above ~100 psig. A sizing and cost estimate, not a compressor selection.
• Compressed-Air Discharge-Pressure Setpoint Savings - Prices the cheapest compressed-air conservation measure there is: turning the discharge pressure down. Because compression power rises with the pressure ratio, the isentropic power ratio between two setpoints gives the exact percent saved - which reproduces the DOE rule of about half a percent of compressor energy per psi. Dropping 120 to 105 psig on a 50 kW compressor saves ~7 percent, roughly $2,000 a year, for turning a knob. The reduced setpoint must still hold the minimum tool pressure after system drop, which is why fixing leaks and piping is what unlocks it. An energy-savings estimate, not a metered result.
• ERV/HRV Sensible Effectiveness and Recovered Load - What the recovery core gives back: a balanced ERV/HRV at a rated sensible effectiveness pre-tempers the incoming outdoor air off the leaving exhaust, so the air reaches the coil at T_oa + eps x (T_ra - T_oa) and the plant is relieved of Q = 1.08 x CFM x eps x dT. A 200 cfm unit at 75 percent on a 10 F design day delivers 55 F air and recovers 9,720 Btu/h - three quarters of the ventilation heating load, the number a designer subtracts before selecting equipment. Sensible only, balanced flow, rated (not frosting-derated) effectiveness. A design aid, not the manufacturer's certified data.
• Makeup-Air Unit Tempering Load (Sensible, Latent, Total) - The heater or coil a makeup-air unit needs: when a kitchen hood or process fan exhausts air, IMC 508 requires a roughly equal makeup supply, tempered so it does not dump raw winter air into the space. Sensible load 1.08 x CFM x dT, optional latent 0.68 x CFM x dW (gr/lb), and the gas input at the burner efficiency - a 2,000 cfm MUA lifting 20 F air to 65 F needs 97,200 Btu/h out and a 121,500 Btu/h burner at 80 percent. Sea-level constants, neutral supply, no duct losses. A design aid, not the engineer's stamped design.
• Demand-Controlled Ventilation Rate from a CO2 Setpoint - The outdoor air per person that holds a CO2 setpoint: the steady-state single-zone mass balance C_in = C_oa + N/Q, solved as Q = N / (C_set - C_oa) with a sedentary generation of about 0.0106 cfm per person. A 1,100 ppm setpoint against 400 ppm outdoor air works out to the classic 15.1 cfm/person office rate; tighten to 800 ppm and the DCV control must drive 1.75x the air. Equilibrium airflow in one fully mixed zone - CO2 is an occupancy indicator, and the ASHRAE 62.1 Ventilation Rate Procedure governs the minimum. A controls and commissioning aid.
• Pipe Flow Reynolds Number and Regime - The dimensionless ratio behind every friction-factor decision, which the friction tiles use implicitly but never show: Re = V D / nu sorts pipe flow into laminar (Re < 2,300), transitional, and turbulent (Re > 4,000). 60 F water at 6 ft/s in a 2 in line runs Re 82,600 - firmly turbulent, as nearly all trade piping is, which is why the Hazen-Williams and Colebrook forms apply; slow it to a crawl and Re 275 laminar, where f = 64/Re and loss is linear in velocity. Enter the kinematic viscosity for the fluid and temperature. An engineering aid; the fluid property data govern.
• Hydronic System Flow from Load and Delta-T - The most fundamental hydronic number, which every balancing valve, pump, and pipe size starts from: GPM = Q / (500 dT) from Q = 500 GPM dT for water, or the chilled-water shortcut GPM = 24 tons/dT. A 10-ton coil at a 10 F design delta-T needs 24 GPM; a 100,000 Btu/h boiler at a wide 20 F delta-T only 10 GPM - the delta-T is the lever that shrinks the pump and pipe for the same load. Pure water at the sea-level factor (adjust for glycol); full load on the delta-T. A design aid; the mechanical engineer of record governs.
• Pump Specific Speed and Impeller Type - The index that says which impeller a duty wants: Ns = N sqrt(Q) / H^(3/4) (rpm, gpm at BEP, ft per stage) classifies the wheel - radial below ~2,000 (high head, low flow, the building-service norm), mixed flow ~2,000-4,500, axial above ~4,500. An 1,750 rpm pump at 500 gpm and 100 ft runs Ns 1,237 (radial); drop the head to 25 ft and Ns jumps to 3,500 (mixed flow) - the head-to-flow ratio, not the size, sets the geometry. US dimensional form; the suction specific speed Nss is separate. An engineering aid; the manufacturer's curves govern.
• Pump Suction Specific Speed (Nss) - The suction-side companion to specific speed: Nss = N sqrt(Q) / NPSHr^(3/4) (rpm, gpm at BEP, required NPSH in ft) indexes how hard a pump works its suction. The Hydraulic Institute design guideline caps Nss near 8,500; above it (especially above ~11,000) a pump shows suction-recirculation damage and shorter life. A 1,750 rpm, 2,000 gpm pump with 25 ft NPSHr sits at a safe Nss 7,000, but on a tighter 16 ft NPSHr it jumps to 9,783 - above the limit, nothing changed but the suction margin. A screening index, not an NPSH-margin calc; the manufacturer's curves govern.
• Refrigerant Mass Flow from Capacity and Refrigeration Effect - The mass flow refrigerant-velocity takes as a given, derived: m_dot = Q / (h1 - h4), the cooling load over the refrigeration effect off the P-h diagram. A 5-ton system with a 60 Btu/lb effect circulates 16.7 lb/min (1,000 lb/h) - the flow the compressor pumps and a tech hands to the velocity check. Warm the liquid line so the effect drops to 50 Btu/lb and the same 5 tons needs 20% more mass flow, the penalty of poor subcooling. Enter the enthalpies from the P-h diagram; steady flow. An engineering aid; the refrigerant property data govern.
• Refrigeration COP and Carnot Limit - The fundamental cycle efficiency seer-eer never reads: COP = (h1 - h4)/(h2 - h1), the refrigeration effect over the compressor work, against its Carnot ceiling T_evap/(T_cond - T_evap) in absolute temperature. A cycle with 60 Btu/lb effect and 25 Btu/lb work runs COP 2.40 (EER 8.19); against a Carnot 6.25 for a 40/120 F lift it is 38% of ideal. A smaller lift - a higher evaporator, a cooler condenser - raises the ceiling, the lever for efficiency. Enter the P-h enthalpies; no parasitic loads. An engineering aid; the property data and state points govern.
• Condenser Total Heat of Rejection - The load the condenser or cooling tower must shed, which the refrigeration side never computes: THR = Q_evap (1 + 1/COP), the evaporator load plus the compressor work. A 5-ton system at COP 2.4 rejects 85,000 Btu/h (7.08 tons) - a 1.42 heat-rejection factor, 42% more than it absorbs; drop to COP 1.5 and it climbs to 100,000 Btu/h, so a struggling system overloads its own condenser and drives head pressure higher still. No heat-recovery split or hermetic-motor heat. An engineering aid; the rated heat-of-rejection data govern.
• Walk-In Cooler Heat Load - Box load = transmission (U x area x deltaT) + infiltration + product + internal, times a safety factor. U 0.05, 800 ft^2, 60 F, +3000/5000/1500, 1.10 -> 2,400 transmission, 13,090 Btu/hr (1.09 tons); a 6 in panel (U 0.03) cuts it to 12,034. Size the evaporator for an ~18 hr run. The box-load method and equipment ratings govern.
• Product Pull-Down Load - Heat to bring product to storage: above freezing Q = m cp deltaT; for a freezer add the latent heat of fusion and the frozen sensible. 2000 lb produce, cp 0.9, 80->35 F, 24 hr -> 81,000 Btu, 3,375 Btu/hr; freezing the same to 0 F -> 358,800 Btu (latent dominates). The product property tables govern.
• Evaporator Design TD and Humidity Band - DTD = box temp - saturated suction, which sets the box humidity: <=10 F ~90% RH (produce), 12-16 F ~75-80% (meat), >16 F <70% (frozen). A 35 F box on a 25 F SST -> 10 F TD, ~90% RH; a drying 17 F suction -> 18 F TD, <70% RH. A smaller TD needs more coil. The coil rating at the design TD governs.
• Whole-Building Heat-Loss Coefficient UA - The whole envelope rolled into one number a degree-day estimate and a balance-point calc both start from: UA = sum(A_i/R_i) + 1.08 x CFM (Btu/h-F). A small house of R-17 walls, R-38 ceiling, R-3 windows, and R-19 floor with 50 cfm infiltration sums to 309.7 Btu/h-F - times a 70 F design difference, a 21,700 Btu/h load. Air-seal to 25 cfm and upgrade the glass to R-5 and the load drops 17%, the two biggest levers a roll-up exposes. Sensible only; enter the natural infiltration cfm. An energy-audit aid, not a stamped Manual J.
• Annual Heating Energy and Fuel Cost from Degree-Days - The annual heating bill an upgrade is measured against: Q = 24 x HDD x UA turns the building UA and the climate's heating degree-days into delivered energy, then fuel = Q/efficiency and cost = fuel x price. A UA 500 house in a 5,000-HDD climate uses 60 MMBtu; at 80% gas that is 750 therms (~$900/yr), or at COP 3.0 electric 5,862 kWh (~$879) - the comparison a fuel-switching decision turns on. Energy scales directly with UA, so a 20% envelope cut is a 20% lower bill. Base-65 steady-state, no gains. An estimate, not a calibrated model.
• Wall Condensation Plane Temperature vs Dew Point - The hidden interface an IR reading of the room-side surface never catches: temperature drops across a wall in proportion to R-value, so T_plane = T_in - (R_inside/R_total)(T_in - T_out), and condensation forms wherever that plane sits at or below the indoor dew point. R-13.5 to the sheathing behind R-4 of cladding on a 70/20 F day puts the sheathing at 31.4 F, 13 degrees below the 44.6 F dew point - a wetting plane; continuous exterior insulation warms it above the dew point, the whole point of the ratio rule. 1-D steady-state screen, no vapor diffusion. A building-science aid, not a hygrothermal analysis.
• Duct Heat Gain/Loss Through Unconditioned Space - Conductive duct heat gain/loss Q = U A dT (U = 1/R) and the resulting supply-air temperature change dT_air = Q/(1.08 CFM). 1,000 CFM of 55 F air in R-4 duct, 100 ft^2, a 120 F attic warms 1.5 F; R-8 halves it. Steady-state conduction, no leakage. A design aid.
• Grille/Register Face Velocity and Free-Area Sizing - Face velocity V = CFM/(gross area x free-area ratio), or the required gross size from a target velocity. 400 CFM to a 500 fpm supply -> 154 in^2 (a 12x14); returns run slower so they size larger. The manufacturer free-area and throw data govern.
• ADPI Room Air Diffusion Selection (ASHRAE) - Whether a supply outlet will actually mix the room, by the ASHRAE Air Diffusion Performance Index: the ratio of the catalog throw T to the room's characteristic length L predicts the ADPI, and each outlet type and cooling load has a T/L for maximum ADPI and a band over which ADPI stays above the comfort threshold. A round ceiling diffuser peaks at T/L = 0.8 (ADPI 88 at a 40 Btu/hr-ft^2 load); a high sidewall grille wants ~1.8 but a heavy 80 Btu/hr-ft^2 load caps it at ADPI 68 no matter the throw - the load, not the outlet, sets the ceiling. Reproduces the ASHRAE Fundamentals selection table (Miller/Nevins, throw per Std 70, ADPI per Std 113). A selection aid; the manufacturer's data and the design engineer govern.
• Vibration Isolation Efficiency (ASHRAE) - Whether the isolators under a fan, pump, RTU, or chiller actually cut the vibration: the isolated system's natural frequency fn = 3.13/sqrt(static deflection in inches) Hz, the disturbing frequency = rpm/60, and the transmissibility T = 1/|(f/fn)^2 - 1| gives the isolation efficiency (1 - T). A 900 rpm fan on 1 in isolators runs fn 3.13 Hz, ratio 4.8, 95% efficient. The trap the tile flags: isolation needs a frequency ratio over sqrt(2) = 1.414 - a slow 200 rpm unit on the same soft mount sits near resonance and AMPLIFIES the shaking 7x, so the fix is a stiffer isolator. Undamped ASHRAE single-DOF idealization; the rated isolator deflection and the mechanical engineer govern.
• Isolator Static Deflection for a Target Isolation - The mount-selection inverse of the vibration-isolation tile: enter the running speed and the isolation efficiency you want, get the static deflection to specify. From T = 1 - efficiency, ratio = sqrt(1 + 1/T), fn = (rpm/60)/ratio, and deflection = (3.13/fn)^2 in. 90% isolation of a 900 rpm fan needs 0.48 in of deflection (a 4.5 Hz mount); tighten to 95% and it nearly doubles to 0.91 in - the softer the mount, the better the isolation, and this says how soft. Feeds straight back into the forward tile. Undamped single-DOF; the isolator selection and the mechanical engineer govern.
• Air Density Correction for Altitude and Temperature (ACFM/SCFM) - Density factor DF = (1 - 6.73e-6 elev)^5.258 x 530/(460+T), converting ACFM to SCFM = ACFM x DF and correcting the 1.08 sensible constant and the rated fan static. 5,000 ft is ~16% thinner; 120 F rooftop air ~9% thinner at sea level. A correction factor; the fan curve governs.
• Moist Air Enthalpy (ASHRAE Psychrometrics) - Total heat content h = 0.240 t + W (1061 + 0.444 t) Btu per lb dry air, the state property every coil and air-mix calculation starts from. Return air at 80 F, W 0.0112 -> 31.48 Btu/lb; 55 F supply, W 0.009 -> 22.97. Pair with outdoor-air-mix for W. A design aid; the chart and equipment ratings govern.
• Cooling Coil Total Load from Enthalpy Difference - Q = 4.5 x CFM x (h_ent - h_lvg) Btu/hr, the full sensible-plus-latent heat a coil removes, from the moist-air enthalpy drop. 2,000 CFM across a 31.48 -> 22.97 Btu/lb drop -> 76,590 Btu/hr (6.38 tons), well above the dry-bulb-only estimate. A design aid; the coil rating governs.
• Coil Bypass Factor and Apparatus Dew Point - Bypass factor BF = (t_lvg - t_adp)/(t_ent - t_adp), contact factor CF = 1 - BF, off the apparatus dew point (coil-surface temperature). 80 -> 55 F off a 50 F ADP -> BF 0.167, CF 0.833; a shallower coil bypasses more and dehumidifies worse. A design aid; the coil rating governs.
• Fan Affinity Laws (Speed / Diameter Change) - For a fixed fan changing speed: airflow scales with speed (Q2 = Q1 r), static with the square (SP2 = SP1 r^2), power with the cube (BHP2 = BHP1 r^3). 10,000 CFM / 1.0 in wg / 5.0 BHP from 900 to 1200 rpm -> 13,333 CFM, 1.78 in wg, 11.85 BHP; slowing to r 0.75 cuts power 58%, the VFD payback in one line. The fan curve governs.
• Pitot Traverse Airflow (Velocity Pressure to CFM) - Field airflow from a Pitot traverse: V = 4005 x sqrt(VP_avg) for standard air, then CFM = V x duct area. 0.15 in wc in a 24x12 duct -> 1551 fpm, 2.0 ft^2, 3102 CFM; a lighter 0.05 in wc reads 896 fpm (the square-root relation). A field measurement, not a substitute for a calibrated flow station.
• Measured Outside-Air Percent from Mixed-Air Temperatures - %OA = 100 (T_ra - T_ma) / (T_ra - T_oa), the field check of the outside-air damper against the design minimum. RA 75, MA 68, OA 40 F -> 20% OA; the damper opening to MA 63 reads 34%. A mixed-air temp outside the return/outdoor band flags a sensor fault; a spread under 10 F warns of low reliability. A field aid.
• Darcy Friction Factor (Swamee-Jain / Colebrook) - The Darcy friction factor for pipe/duct head loss: laminar f = 64/Re (Re < 2300), turbulent via the Swamee-Jain fit to Colebrook, f = 0.25/[log10(eps/D/3.7 + 5.74/Re^0.9)]^2. Re 100,000, eps/D 0.0003 -> f 0.0195 (the Moody value for commercial steel); Re 1500 -> 64/1500 = 0.0427. Feeds h = f (L/D) V^2/2g. A design aid.
• Condensate Rate and Drain Size - The gallons per hour a coil sheds from the tonnage, the minimum drain-line size by IMC 307.2.2, and the fall over a run at the code slope. The per-ton rate is an editable field estimate, since condensate tracks the latent load.
• Recovery-Cylinder 80% Fill - The maximum net refrigerant a recovery cylinder may legally hold under the 80% fill rule, from the stamped water capacity and the refrigerant's liquid density, plus the headroom left and the fill / do-not-fill action.
• HVAC Equipment Circuit (MCA / MOCP) - The minimum circuit ampacity and maximum overcurrent device for a condenser or heat pump per NEC 440.33 / 440.22 -- MCA is 125% of the compressor RLA plus the other loads, MOCP is 175% to the next standard size down (225% to start) -- with a check of the installed breaker against the nameplate maximum.
• Run Capacitor Microfarad Check - The in-circuit capacitance of a motor run capacitor from the measured volts and amps (C = I / (2 x pi x f x V) at 60 Hz, the ~2652 x amps / volts field rule) compared to the nameplate rating and an editable +/-6% tolerance band, with a good / weak / replace verdict.
• Vacuum Decay (Blank-Off) Test - The evacuation standing-decay verdict: from the micron level at valve-off, the level after a timed hold, and the hold time, the total rise and the rise rate, with a tight-and-dry / moisture / leak reading against an editable 500-micron pass ceiling (ACCA Standard 4 / AHRI field practice).
• Nitrogen Pressure Test (Temperature-Corrected) - Separates a real leak from the thermal swing in a standing nitrogen pressure test using Gay-Lussac's law (P/T constant at fixed volume, absolute units): the temperature-corrected expected gauge pressure, the pressure lost below it to a leak, and a holds / leak verdict against an editable tolerance.
• Gas-Meter Clocking (Actual Firing Rate) - The actual firing rate of a gas appliance by clocking the meter: from the seconds for one revolution of a known test dial, the dial size, and the fuel heating value (default 1030 BTU/cf natural gas, ~2500 LP), the gas flow in cfh and the input in BTU/hr, with a firing-on-rate / overfired / underfired verdict against the nameplate. Clock with every other gas appliance off.
• Furnace Temperature Rise and Derived Airflow - The supply-minus-return temperature rise checked against the rating-plate range (default 40 to 70 F), and the airflow derived from the heat output via Qs = 1.08 x CFM x delta-T: from return / supply temperatures, furnace input, and efficiency, the rise, the output, and the CFM, with a rise-low / in-range / rise-high reading. The rating plate governs.
• Blower-Door Air-Tightness (ACH50, Natural Infiltration, Code Check) - Turns a blower-door reading into the number every weatherization job turns on: ACH50 = CFM50 x 60 / conditioned volume, checked against the IECC R402.4.1.2 limit (<= 3 ACH50 in climate zones 3-8, <= 5 in zones 1-2), plus the year-round natural infiltration (the LBL divide-by-N rule, in air changes and cfm) that the ASHRAE 62.2 ventilation sizing and the infiltration load both consume. ASTM E779 / E1827 are the test methods; the N-factor varies with climate, height, and shielding. A field normalization, not a rater's signed test report.
• ASHRAE 62.2 Whole-House Mechanical Ventilation Rate - The residential whole-house ventilation a tight house needs: Qtot = 0.03 x conditioned floor area + 7.5 x (bedrooms + 1) (ASHRAE 62.2-2019), and the continuous fan the installer must provide, Qtot minus the infiltration credit the envelope already leaks. The tighter the house tests, the smaller the credit and the larger the fan 62.2 requires - the trade-off a blower-door retest makes concrete. The conservative default is zero credit. Local kitchen/bath exhaust is separate. A sizing aid, not a 62.2 compliance certificate.
• Infiltration Heating / Cooling Load (Sensible + Latent) - The load the air the envelope leaks adds to the design day: sensible Qs = 1.08 x cfm x delta-T and latent Ql = 0.68 x cfm x delta-grains, from the natural infiltration cfm a blower-door test produces. In cooling the latent half (the moisture the leaking air drags in) is nearly as large as the sensible, which is why air-sealing pays back on both design days. The 1.08 and 0.68 are sea-level standard-air constants. The infiltration component only, not a stamped Manual J.